The invention relates to a shaft seal for rotating shafts in turbo-machines or other pressurized machine. In particular, the present invention, in common with a known form of shaft seal, provides a shaft seal comprising a sealing element, a rotary sealing part mounted coaxially with the sealing element and forming therewith a contactless primary seal between opposed faces of the sealing element and rotary sealing part to substantially prevent fluid flow across the primary seal from a high pressure radial side to a low-pressure radial side, a seal housing, an annular pusher disc disposed about a forwardly extending sleeve portion of the seal housing and coaxially with the sealing element, biasing means acting on the pusher disc to urge the sealing element axially towards the rotary sealing part, and a first sealing member disposed between the pusher disc and the forwardly extending sleeve portion in communication with the high-pressure radial side to provide a secondary seal between the high-pressure and low-pressure radial sides.
Non-contacting shaft seals are often used with machinery for the compression or expansion of gas (hydrogen, natural gas, air, etc.) where the transmission of gas along the shaft needs to be prevented. Due to the high-pressure, high-speed machinery which is normally used, the shaft seals are chosen to be of non-contact type, in order to reduce heat build up in the seals and the wear of the sealing parts and/or in order to avoid the complexity of oil seals and their associated systems.
Non-contacting operation avoids this undesirable face contact when the shaft is rotating above a certain minimum speed, which is often called a lift-off speed.
Non-contacting shaft seals provide advantages over seals where the sealing surfaces contact one another, due to reduction in wear and the lower heat generation. Articles entitled “Fundamentals of Spiral Groove Non-contacting Face Seals” by Gabriel, Ralph P. (Journal of American Society of Lubrication Engineers Volume 35, 7, pages 367-375), and “Improved Performance of Film-Riding Gas Seals Through Enhancement of Hydrodynamic Effects” by Sedy, Joseph (Transaction of the American Society of Lubrication Engineers, Volume 23, 1 pages 35-44) describe non-contacting seal technology and design criteria and are incorporated herein by reference.
As with ordinary mechanical seals, a non-contacting face seal consists of two principal sealing elements. At least one of the sealing elements is provided with shallow surface recesses.
These recesses are taper-shaped perpendicular to and concentric with the axis of rotation, the tapering being in the direction opposite to the direction of rotation of the shaft. In known contactless face seals, both sealing elements, in the form of rings, are positioned adjacent to each other with the sealing surfaces in contact at conditions of zero pressure differential and zero speed of rotation. One of the rings is normally fixed to the rotatable shaft by means of a shaft sleeve, the other being located within the seal housing structure and allowed to move axially. The shaft seal is designed to enable axial movement of the sealing ring and yet prevent or substantially minimize leakage of the sealed fluid. For this reason, a sealing member is placed between the ring and the housing.
As mentioned above, to achieve non-contacting operation of the seal, one of the two sealing surfaces is provided with shallow surface recesses, which act to generate pressure fields that force the two sealing surfaces apart. When the magnitude of the forces resulting from these pressure fields is large enough to overcome the forces that urge the seal faces closed, the sealing surfaces will separate and form a clearance, resulting in non-contacting operation.
As explained in detail in the above-referenced articles, the character of the separation forces is such that their magnitude decreases with the increase of face separation. Opposing or closing forces, on the other hand, depend on sealed pressure level and as such are independent of face separation. They result from the sealed pressure and the spring force acting on the back surface of the axially movable sealing ring. Since the separation or opening force depends on the separation distance between sealing surfaces, during the operation of the seal or on imposition of sufficient pressure, differential equilibrium separation between both surfaces will establish itself. This occurs when closing and opening forces are in equilibrium and equal to each other. Equilibrium separation constantly changes within the range of gaps. The goal is to have the low limit of this range above zero. Another goal is to make this range as narrow as possible, because on its high end the separation between the faces will lead to increased seal leakage. Since non-contacting seals operate by definition with a clearance between sealing surfaces, their leakage will be higher than that of a contacting seal of similar geometry. Yet, the absence of contact will mean zero wear on the sealing surfaces and therefore a relatively low amount of heat generated between them. It is this low generated heat and lack of wear that enables the application of non-contacting seals to high-speed turbo machinery and other pressure machines, where the sealed fluid is gas. Turbo compressors are used to compress this fluid and since gas has a relatively low mass, they normally operate at very high speeds and with a number of compression stages in series.
As explained in the above-referenced articles, the effectiveness of the seal is largely dependent upon the so-called balance diameter of the seal. This is also true for contact seals.
When pressure is applied from the outside diameter of the seal, reduction of the balance diameter results in a greater force pushing the two sealing faces together and so a smaller gap between the faces. Thus, less gas is leaked from the system.
Known compressors have been used for compressing gas at inlet pressures of some 200 bar to delivery pressures of some 500 bar. Contactless shaft seals of the kind described above are typically used to seal against the compressor inlet pressure. The trend in compressor requirements nowadays is towards higher inlet and delivery pressures. However, such pressure levels give rise to a problem with the contactless shaft seals described above, as is now explained with reference to FIGS. 1, 1a. 
FIG. 1 is a partial longitudinal sectional view through the shaft seal showing the relevant structural elements of a known non-contacting shaft seal of the type described above. The shaft seal is incorporated in a turbo-machine (not shown), such as a compressor in this example. There is shown a shaft seal 1 having a (non-rotating) sealing element or ring 2 mounted coaxially with the shaft axis (denoted by reference numeral 3), and a rotary sealing part or ring 4 located coaxially with the sealing ring 2, and therefore also with the shaft axis 3. It will be appreciated that the vertical sectional view of FIG. 1, for simplicity, shows only the portion of the shaft seal located above the shaft axis. The sealing ring 4 is mounted on an inner sleeve 5 having a radial flange 5a against which the sealing ring 4 abuts, the sleeve 5 being mounted on the shaft 6 such that the shaft 6, inner sleeve 5 and rotary sealing ring 4 co-rotate as a single rotary element. In addition, a locating sleeve 7 is bolted to inner sleeve 5. The assembly comprising components 4, 5 and 7 is prevented from displacement in one axial direction by a locating ring 21 and in the opposite axial direction by the high pressure acting inside the compressor.
The shaft seal also has a seal housing 8 and an annular pusher disc 9 disposed between a radially inward flange portion 8b of the seal housing 8 and sealing ring 2 and loosely fitted around a forwardly extending sleeve portion 8a of the seal housing. A plurality of biasing springs (one of which, 10, is shown in FIG. 1), located at the same axial position in respective blind holes 11 in radially inward flange 8b and distributed about the shaft axis, act against the pusher disc 9 to urge it against the sealing ring 2. The (non-rotary) sealing ring 2 and rotary sealing ring 4 together form a contactless primary seal when the turbo-machine (or pressurized machine) is in operation, which substantially prevents fluid flow between the sealing faces of the primary seal, from the high pressure radially outer side to the low pressure radially inner side. The sealing face of sealing ring 2 has shallow grooves cut into its front surface to generate the required separation between the sealing faces of sealing rings 2, 4. Alternatively, the grooves could be formed in the rotary sealing ring 4.
Preferred designs for the grooves are given in more detail in Publish International Application WO-A-96/15397 of Dresser-Rand Company and the preferred designs for the groove are incorporated herein by reference. The sealing element 2 is normally made from carbon or other suitable material.
As shown in FIG. 1, the sealing element 2 is afforded limited axial movement against the biasing force of the springs 10. These springs provide a relatively small net biasing force so that when the shaft is rotating at normal speed, the generated separating forces cause the sealing ring 4 to separate from the sealing ring 2. The gap between these rings adjusts itself such that the generated opening forces on the one hand and the sum of the generated closing forces and the spring biasing force on the other hand are equal to one another. However, when the shaft is at rest the springs act to move the sealing ring 2 into contact with the rotary sealing ring 4.
A high-pressure gas is supplied to the radially outer edge of the seal rings 2, 4. Normally, this gas would be derived from the working fluid of the machine. However, it could instead be a clean gas suitable for venting into the atmosphere. In that event, the vented gas can be a combustible gas which is piped to burn (flare).
The high pressure at the high-pressure radial side acts around the rear face of sealing element 2 down to a so-called equilibrium balance diameter. Located in a stepped recess 14 formed in the front face of the pusher disc 9 adjacent its inner circumference is a secondary seal 12 which seals against both the seal ring 2 and the forwardly extending sleeve portion 8a of housing 8. This secondary seal serves to prevent the high pressure venting around the rear face of sealing element 2 or behind the pusher disc 9 to the low-pressure radial side (atmospheric pressure). The balance diameter is determined essentially by the contact line of secondary seal 12 with the forwardly extending sleeve portion 8a of housing 8. The secondary seal 12 can be of any suitable form, such as a conventional O-ring, as shown, or a spring-energised U-seal. Other forms of seal are possible and the precise form selected is not material.
In use of the shaft seal 1, the high-pressure working fluid of the compressor is admitted to the high-pressure radial side of the primary seal. This pressure acts on the front face of the pusher disc 9 down to the circular line of sealing of the secondary seal 12 against the sealing ring 2. The high-pressure fluid also acts against the rear face of pusher disc 9 down to the balance diameter. The secondary seal 12 seals the applied high-pressure from the low-pressure radial side, which is at atmospheric pressure where a single shaft seal is used or, if multiple shaft seals are provided in cascade, at a lower pressure than the pressure to be sealed. Because of the pressure differential acting on the rear face of pusher disc 9 down to the balance diameter, there is a net closing force (to the left in FIG. 1) acting on the pusher disc 9 against the sealing ring 2 at all times. This closing force is supplemented by the action of the biasing springs 10, and these closing forces are applied in the closing direction against sealing ring 2. In addition, the high pressure fluid acting on the front faces of sealing ring 2 produces an opening force, while the high pressure fluid acting on the rear faces down to the sealing diameter of secondary seal 12 produces a closing force. Still further, the taper-shaped surface recesses or grooves cut in the front face of sealing ring 2 (or rear face of sealing ring 4) generate separating pressure fields acting between the sealing rings 2, 4, the magnitude of the pressure fields depending on the rotational speed of the compressor shaft. The high pressure to be sealed, the depths of the recesses or grooves and the size of the gap between the sealing rings 2, 4 also influence the magnitude of the pressure fields. Whether the sealing rings 2, 4 of the shaft seal are in contact or separated depends on the magnitudes of the generated opening and closing forces, and the net spring biasing force.
When the compressor is started up, as the rotational speed of the shaft 6 initially starts to build up, the primary seal maintains a substantially fluid-tight seal between the high-pressure and low-pressure radial sides, by virtue of sealing contact between the sealing rings 2, 4. Under these conditions, the net separating force generated by the primary seal is insufficient to overcome the sum of the spring biasing forces and the net closing force acting on the primary seal due to the applied high-pressure.
However, when the compressor shaft speed reaches a sufficient value such that the applied fluid pressure is adequate to generate a separating force that overcomes the net closing force acting on the sealing ring 2, this sealing ring will start to move away from the sealing ring 4 into an equilibrium position in which it maintains a contactless seal between the rotating sealing ring 2 and the non-rotating sealing ring 4. As described above, the secondary seal 12 functions at all times to prevent leakage of high-pressure fluid past the rear face of sealing ring 2 and the pusher disc 9.
Shaft seals of the type described above with reference to FIG. 1 operate satisfactorily at typical sealing pressures of compressors that have been manufactured in the past. Typically, such compressors have been manufactured for compressing gases at pressures of typically from about 200 bar to about 500 bar. However, the industry is now demanding compressors to compress gas from 300 bar or more to 800 bar or more. On the other hand, it has been found that existing shaft seal designs are not adequate to withstand such inlet-pressure values, for the reasons now to be described with reference to FIG. 1a. 
This Figure shows, in deliberately exaggerated manner for the purposes of illustration, the effect of operating under such high-pressure values. As shown in the Figure, the high-pressure acting on the outer face of the forwardly extending sleeve portion 8a of the housing 8 between the seal 12 and the junction with the flange portion 8b deforms the flange portion inwardly with a deflection increasing with increasing axial distance in the axial direction away from the flange portion. This torsional deformation is indicated by letter A in FIG. 1a. Correspondingly, the high pressure acting against the inside (front) face of radial flange portion 8b torsionally deforms that flange rearwardly, as indicated by arrow B. The consequence is that, as shown in FIGS. 1a, 2a, the very small gap normally existing between the inner face of the sealing ring 2 and the outer face of the forwardly extending sleeve portion 8a of the housing 8 is enlarged. With increasing high-pressure acting against the secondary seal 12 and widening of the gap between the sealing ring 2 and the forwardly extending sleeve portion 8a, a bead (not shown) starts to form as the secondary seal 12 starts to be extruded through the widening gap. When there is no such bead on the secondary seal 12, this seal offers little frictional resistance to the rearward axial sliding of the pusher disc 9. However, when the bead (not shown) starts to form, the frictional resistance increases, potentially significantly and even to the point where the pusher disc can become united with the forwardly extending sleeve portion 8a. Furthermore, as the bead (not shown) continues to grow, an increasingly unstable situation can develop whereby the sealing ability of the secondary seal 12 is progressively lessened due to the continuing extrusion, until eventually an unstable situation is reached in which the seal 12 is expelled or blown out through the gap, resulting in failure of the shaft seal. It is noted that the bead (not shown) does not normally form around the entire rear circumferential region of the secondary seal 12 but generally only at a single angular position about the seal circumference.
One possible solution to this problem that has been considered is to minimise the gap existing between the sealing ring 2 and the pusher disc 9 when the shaft seal is not in use, but there is a limit to how much this gap can be reduced because the pusher disc 9 must be free to undergo limited axial movement when the shaft seal is not in operation. Furthermore, radially inward deflection of the sleeve portion 8a is inevitable, yet this sleeve must not be allowed to come into contact with the (rotating) shaft inner sleeve 7 under full operating pressure.
Another potential solution which has been considered is to use harder materials for forming the sealing parts of the secondary seal 12. However, there is a limit to how hard the selected materials can be, particularly since harder materials are less effective to provide the required sealing effect and they also increase the friction forces generated.
Spring energised polymer seals have been proposed. However, the operating pressure at which beads start to form on such seals is about 200-250 bar.